The invention relates to an impeller for pumps, in particular for radial pumps, and to a pump, in particular a radial pump with an impeller of this kind, and also to a method for operating a pump of this kind.
In centrifugal pumps with one or more impellers, the resultant of all the axial forces acting on the impeller or the impellers during operation can reach considerable values. Without additional measures this resultant force which is termed axial thrust, would be transferred via the pump shaft to the bearing and would correspondingly put a heavy load on this. Suitable constructional measures are known from the prior art for reducing the axial thrust, for example by means of a dual-flow pump arrangement with a mirror symmetrical design of the impellers, or in single flow pumps by providing sealing gaps on both sides of the impeller and openings in the impeller which connect the suction side with the reverse side of the impeller. A further possibility is to relieve the pump of axial pressure by means of a suitable relief device, such as a relief ring or piston, for example.
For the characterisation of a centrifugal pump impeller the specific rotational speed nq is often used which is calculated in known manner from the capacity or pump flow Q, head H and rotational speed n. In pumps for small specific rotational speeds nq, of for example smaller than 15 min−1, corresponding to a comparatively large head of greater than 50, 75 or 150 m and comparatively small capacities of smaller than 100, 50 or 25 m3/h for example, in particular in those kinds of process pumps, the problem arises that the above-mentioned measures for the axial thrust compensation can only be applied to a restricted extent. Thus a dual flow version of a pump for small capacities requires a considerable amount of additional cost and complexity in comparison with a single flow design. A balance piston is likewise comparatively costly and for this reason is mainly used in larger multi-stage pumps. For a single stage process pump, sealing gaps are normally provided on both sides of the impeller. An axial thrust compensation by means of sealing gaps on both sides of the impeller and openings in the impeller is, however, only possible when the impellers are closed. In the case of closed impellers with small specific rotational speeds nq there is the problem of the manufacture, since the outlet widths of impellers of this kind are in the region of few mms and the manufacture of closed impellers with small outlet widths is difficult and expensive from the point of view of casting technology.
A further disadvantage of closed impellers is the high impeller friction losses (also called impeller side friction losses) and clearance gap losses which impellers of this kind have at small specific rotational speeds nq. At a specific rotational speed nq, of 8 min.−1 for example, the impeller friction loss alone amounts to 30% or more. Closed impellers for small specific rotational speeds nq thus show a comparatively low degree of efficiency.
For the above-mentioned reasons, impellers for specific rotational speeds nq smaller than 10 or 15 min.−1 are often designed to be half-open. This has advantages from the point of view of casting technology and the wheel friction of half-open impellers is considerably less than those of closed impellers. Half-open impellers for specific rotational speeds nq less than 10 or 15 min.−1 have the disadvantage however that the axial thrust compensation is difficult and the impeller friction losses are still very high.
An additional problem of closed and half-open impellers for small specific rotational speeds nq is the tendency to instability in the part load region, i.e. impellers of this kind have a characteristic curve which is either unstable (corresponding to a falling characteristic curve, if the pump flow Q approaches 0) or only just stable (corresponding to a characteristic without a notable rise when Q approaches 0).